1. Field of the Invention
The present invention relates to a gas seal, and more specifically, to a gas seal with spiral grooves.
2. Description of the Related Art
Dry-running gas lubricated seals provide significant economic benefits over the prior art oil seals which consume significantly more power due to the greater viscosity of oil as compared to gas. Furthermore gas seals do not require expensive apparatus for cleaning and cooling the oil.
One of the first working concepts for a combination hydrostatic and hydrodynamic gas seal was discussed in U.S. Pat. No. 3,499,653 to Gardner. This seal has a spiral groove pattern on one of the sealing faces which allows gas pressure to provide hydrostatic separation and rotation to provide hydrodynamic separation of the sealing faces. U.S. Pat. Nos. 4,212,475 to Sedy and U.S. Pat. No. 4,768,790 to Netzel et al. discuss improvements and refinements to spiral groove gas seals.
However, dry-running gas seals exhibit some areas of performance which are less than optimal. First, improvement of the dynamic tracking of the seal faces during rotation is needed to allow the seals to ride on a thinner gas film and track the runout on the face of the rotating ring. Runout on the faces of rotating rings is typically 0.0007 to 0.0030 inch T.I.R., which is considerably greater than the 100 to 300 micro inch gas film on which the seal faces ride. Therefore, it is critical for the seal faces to be able to take on a swash plate action (i.e., a simultaneous oscillation about two axes in the sealing face) in order to follow the total runout on the rotating ring. It is this ability to takeon the swash plate action that is called dynamic tracking. The dynamic tracking capabilities of the seal faces are affected by the drag imparted on the seal ring by "O" ring forces and by the mass of the axially mobile stationary seal structure.
In a conventional gas seal, a "O" ring is placed between the stationary seal ring and the stationary seal ring carrier, and the seal ring is typically manufactured from materials having low thermal expansion coefficients, such as carbon graphite, silicon carbide and tungsten carbide. The ring carrier in such a conventional gas seal is fabricated from metal, such as stainless steel, Inconel 625 or another metal having good corrosion resistance, temperature resistance and the required strength. These metals from which the carriers are formed have coefficients of expansion approximately three times that of the typical seal ring materials.
The seal rings are mounted within a stationary housing and mounted over a metal balance diameter which serves to center the seal ring with a clearance used to compensate for differential thermal growth between the seal ring and the balance diameter. For a seal with a 4.625 inch balance diameter, a diametrical clearance of 0.015 to 0.019 inches would be common. During operation, the seal assembly becomes hot, with temperatures up to 350.degree. to 400.degree. F. not uncommon. When the assembly heats, the balance diameter expands relative to the seal ring and then centers the seal ring.
However, differential thermal expansion between the balance diameter and the seal ring will create variability in the squeeze force on the "O" ring. This variability will adversely effect the dynamic tracking of the stationary seal ring, since the stiffness (i.e., the force preventing opening and closing) of the sealing faces imparted by the "O" ring may be as high as 100,000 pounds per inch.
Attempts have been made to eliminate the differential thermal squeeze on the "O" ring by placing it between a metallic "O" ring holder and the metallic carrier. However, these devices, such as the one in Netzel et al., have significant mass, and therefore their inertia might negatively affect the dynamic tracking capability of the sealing face.
A further problem of dynamic tracking is associated with placement of the seal ring carrier on a high speed shaft. Typically, an alloy such as 4340 is used for the shaft itself which has a different thermal expansion coefficient than the rotary ring carrier. Although not as great as the differential thermal expansion between the stationary seal ring and the carrier, the loosening of the rotary seal ring carrier on the shaft is of greater concern due to the high speed of rotation. Typically, for a 4.00 inch shaft of 4340 with a stainless steel rotary ring carrier, approximately 0.003 inches loosening at 400.degree. F. results from differential thermal expansion. At a typical rotating shaft speed of 16,000 rpm, a ten pound carrier creates centrifugal unbalanced force of 218 pounds.
Additional loosening of the carrier on the shaft occurs due to differential centrifugal expansion. It is well known that a hollow sleeve grows centrifugally at a much greater rate than a solid shaft. This factor gives an additional degree of loosening at high speed. For the same 4.00 inch shaft at 16,000 rpm, an additional loosening of 0.0015 to 0.002 inches would be experienced.
In the art, shrink fits have often been used to secure seal carriers on a rotating shaft in order to eliminate the loosening due to thermal and centrifugal effects. However, the shrink fit makes installation and removal of seal assemblies very difficult in many cases.
Another area where improvement in gas seal performance is desirable concerns gas film thickness. Conventional spiral groove gas seals tend to exhibit a substantial difference in gas film thickness between hydrostatic (i.e., non-rotating) and hydrodynamic (rotating) operation due to the significant pumping force provided by the spiral grooves. The larger film thickness at high rotations allows for greater leakage across the sealing faces than would otherwise be desired.
A final area of sub-optimal performance in current gas seal design relates to the safety problems present in conventional gas seals. A conventional seal holds the rotary seal ring with pins passing from the rotary seal ring carrier into the seal ring. The pins can cause significant stress concentration in the rotary seal ring and lead to fracture of the ring. Furthermore, the problem is accentuated in the conventional gas seal design by the lack of any retaining structure for the rotary seal ring should catastrophic failure occur. At operating speeds up to 16,000 rpm, a fractured ring can cause significant damage to the machine in which it is operating as well as expensive down time in repair.